1. Field of the Invention
The present invention relates to a hydraulic power steering apparatus suitable for use in vehicles and the like. More particularly, the present invention relates to a hydraulic power steering apparatus which is capable of reducing the power consumption of a vehicle engine by decreasing the flow rate supplied from a pump to a control valve at a low load pressure. Further, the present invention relates to an improvement in a hydraulic power steering apparatus described in the co-pending U.S. application Ser. No. 08/304,923 of the same assignee as this application and another assignee.
2. Discussion of the Related Art
A hydraulic power steering apparatus is usually provided with a hydraulic pump and a flow control valve for supplying a pressurized fluid to an assist force generating mechanism at a constant flow rate. In such a power steering system, the energy consumed by the hydraulic pump increases in accordance with an increase in the flow rate of the pressurized fluid. Therefore, the conventional hydraulic pump consumes a large amount of energy or power.
In order to solve the above-mentioned problem, improved power steering apparatuses have been proposed for reducing the flow rate of pressurized fluid during high speed traveling. Examples of such power steering apparatuses are shown in Japanese Patent Publication No. 54-5571 and U.S. Pat. No. 4,714,413. For example, in the apparatus of Japanese patent publication No. 54-5571, there is provided a vehicle speed sensor S, an amplifier A for amplifying a signal from the sensor S, and an electromagnetic valve SV which responses to the amplified speed signal, as shown in FIG. 1. The valve SV operates to reduce the pressure in a spring chamber of a flow control valve FC in accordance with an increase of the vehicle speed, thereby reducing the flow rate of the pressurized fluid supplied to the assist force generating mechanism which is composed of a rotary valve RV and a power cylinder PC. This system reduces the energy consumption of the hydraulic pump p. The power steering system also has the desirable characteristic that assist forces generated during high speed traveling are smaller than those during low speed traveling.
The conventional power steering apparatuses, however, have the following drawback. Namely, when the steering wheel HD is turned, the flow rate of the fluid flowing through the electromagnetic valve SV increases compared with that when the steering wheel HD remains at its neutral position, because the opening degree of the electromagnetic valve SV depends exclusively on the vehicle speed. Thus, the flow rate of the fluid to the rotary valve RV is decreased, whereby the characteristics of the power assist during high speed is undesirably changed. To avoid this problem, there must be provided a pressure compensation valve which acts upon an increase of the pressure upstream of the rotary valve RV.
To improve this drawback, another power steering apparatus has been proposed, which, as shown in FIG. 2, is mainly composed of an engine-driven pump 100 for discharging operating fluid, a reservoir 101, a power cylinder 102 for assisting the steering operation, a control valve 103 for controlling operating fluid which is supplied from the pump 100 to the power cylinder 102, upon rotation of the steering wheel (not shown), a flow control valve 108 and a load pressure responsive valve 111.
The flow control valve 108 has at the back thereof a spring chamber 110 in which a spring 107 is disposed. The flow control valve 108 is disposed in a bypass passage 106 to control the flow of fluid flowing from an inlet port to an outlet port of the flow control valve 108. A port of the spring chamber 110 is connected to a supply passage 104 via a control orifice 109 and to the reservoir 101 via a relief valve (not numbered). The flow control valve 108 responds to the pressure difference across a metering orifice 105 disposed in the supply passage 104 which connects the pump 100 to the control valve 103, so that the bypass passage 106 is opened and closed by the flow control valve 108 to maintain the flow rate of operating fluid supplied to the control valve 103 constant.
The port of the spring chamber 110 is also connected to the reservoir 101 via the load pressure responsive valve 111. A control spool 112 of this valve 111 is directly slidably inserted in a pump housing 100a. A variable orifice 111A composed of a slits 112a formed at a rear end of the control spool 112 and a annular groove 114 are formed in the pump housing 100a.
When the steering wheel is at a neutral state, the load pressure remains low. Therefore, the control spool 112 of the valve 111 remains urged to the left as viewed in FIG. 1 by a spring 113 arranged at the rear end of the control spool 112, so that it maintains the largest opening area of the variable orifice 111A. With this state, the pressure in the spring chamber 110 of the flow control valve 108 is released to the reservoir 101 via the variable orifice 111A and remains low. This causes the bypass passage 106 of the flow control valve 108 to open much more, so that the operating fluid from the pump 100 is bypassed to the reservoir 101 much more, thereby decreasing the flow rate of the fluid supplied to the control valve 103. As a result, the energy consumed by the pump 100 can be reduced.
When the steering wheel is turned, the pressure on the supply passage 104 upstream of the control valve 103 (that is to say, the "load pressure") gradually increases. When the load pressure exceeds a predetermined pressure in this state, the control spool 112 is moved to the right as viewed in FIG. 2 against the force of the spring 113 to diminish the opening area of the variable orifice 111A. When the load pressure further increases, the opening area of the variable orifice 111a is completely closed. This causes the pressure in the spring chamber 110 of the flow control valve 108 to increase, so that the flow control valve 108 is displaced to close the bypass passage 106. Therefore, the flow rate supplied to the control valve 103 is increased as the load pressure increases, so that the power assist is generated.
The load pressure responsive valve 111, however, has the following drawbacks. Once the spool 112 begins to move against the spring force of the spring 113, it is moved to the right end without taking an intermediate position. This causes an abrupt increase of the power assist, thereby giving the driver an unpleasant feeling.
Another related art is Japanese unexamined patent publication no. 6-171522, which discloses another power steering apparatus in which the power consumption is reduced during low load pressure and high speed traveling. This power steering apparatus has the same configuration as shown in FIG. 2 and is further provided with a traveling speed responsive valve which is arranged in parallel to the load pressure responsive valve 111. The traveling speed responsive valve controls the flow rate bypassing the load pressure responsive valve 111 so as to change the degree of the opening area thereof in response to the vehicle speed. At low traveling speeds, the pressure in the spring chamber 110 is discharged to the reservoir 101 mainly through the variable orifice 111A of the load pressure responsive valve 111, thereby increasing the flow rate supplied to the control valve 103, because a variable orifice of the traveling speed responsive valve remains completely closed. On the other hand, at high traveling speeds, the pressure in the spring chamber 110 is discharged to the reservoir 101 mainly through the variable orifice of the traveling speed responsive valve, thereby reducing the flow rate supplied to the control valve 103, because the traveling speed responsive valve is larger in opening area than the load pressure responsive valve 111. With this operation, the energy consumption and the stability at high speed traveling are ensured.
In such a power steering apparatus, when the vehicle runs at high speeds with the steering wheel at around the neutral position thereof, the opening area of the traveling speed responsive valve is set to be responsive to the traveling speed, so that the pressure in the spring chamber 110 of the flow control valve 108 is drained to the reservoir 101 in dependence upon the vehicle speed. Thus, the flow rate supplied to the control valve 103 can be reduced in response to the traveling speed.
However, in the power steering apparatus constructed above, when the load pressure increases upon rotation of the steering wheel during high speed traveling, the differential pressure across the variable orifice of the traveling speed responsive valve increases to increase the flow rate which is drained to the reservoir 101. As a result of this operation, the flow rate supplied to the control valve 103 cannot be controlled in accordance with the traveling speed. To solve this problem, a power steering apparatus must be provided with a pressure compensation valve downstream of the traveling speed responsive valve. The addition of such a pressure compensation valve causes manufacturing costs to increase.
The assignee of this application has proposed in the cross-referenced co-pending U.S. application Ser. No. 08/304,923 an improved power steering apparatus provided with a bypass control valve 120 as shown in FIG. 3, instead of the load pressure responsive valve 111 in FIG. 2. The bypass control valve 120 has a load pressure inlet port 123A at one end and a pilot port 123B at the other end. The inlet port 123A is connected upstream of the control orifice 109, while the pilot port 123B is connected downstream of the control orifice 109. The valve 120 comprises a control spool 121 having a first end facing the inlet port 123A, a ball 125 held on a second end of the control spool 121 opposite to the first end, a valve seat member 124 disposed adjacent to the pilot port 123B to permit the ball 125 to seat thereon, and a spring 122 disposed between the control spool 121 and the valve seat member 124 to urge the control spool 121 in a direction to separate the ball 125 from the valve seat member 124. The valve seat member 124 is formed with a passage in communication with the pilot port 123B. The ball 125 faces an inner opening of the passage to form a pressure receiving area which is smaller in area than the first end of the control spool 121. A chamber formed between the control spool 121 and the valve sheet member 124 is in communication with the reservoir 101 through a drain port 123C.
When the control valve 103 is in its neutral state, the load pressure is at a low level P.sub.A so that only a small differential pressure is produced across the control orifice 109. In this state, a variable throttle 120A of the bypass control valve 120 is fully opened due to the spring force of the spring 122. As a result, the spring chamber 110 of the flow control valve 108 communicates with the reservoir 101, so that the pressure in the spring chamber 110 is lowered. This causes the flow control valve 108 to retract so as to open the bypass passage 106. Accordingly, a substantial part of the operating fluid discharged from the pump 100 is mostly bypassed to the reservoir 101. With this operation, the flow rate of the operating fluid supplied to the control valve 103 is reduced to the lowest value Q.sub.A, as shown in FIG. 6.
When a steering wheel (not shown) is turned, the upstream pressure of the control valve 103, i.e., load pressure, gradually increases as is well known in the art. When the load pressure increases, the differential pressure across the control orifice 109 increases. When the differential pressure reaches a predetermined level, the control spool 121 is moved toward the pilot port 123B against the spring force of the spring 122, thereby decreasing the opening area of the variable throttle 120A. When the differential pressure across the control orifice 109 is further increased due to a further increase of the load pressure, the control spool 121 of the bypass control valve 120 closes the pilot port 123B, so that the flow rate q of pilot fluid flowing into the pilot port 123B is decreased to zero, as shown in FIG. 5. With this operation, the pressure in the spring chamber 110 of the flow control valve 108 increases, so that the flow control valve 108 moves toward the direction to close the bypass passage 106. As a result, the flow rate of the operating fluid to the control valve 103 is increased as the load pressure increases. When the load pressure reaches P.sub.B, the flow rate reaches the maximum rate Q.sub.B sufficient to generate a required assisting force.
However, the control spool 121 is disposed at the inner surface of the pump housing 100a, so that the annular groove 114 has to be formed in pump housing 100a. This requires machining of the pump housing 100a from the inside thereof. Therefore, the machining of the annular groove 114 is difficult.